There is known a hydraulic-pump driving system controlling device as shown in FIG. 7, which is a conventional over-loading prevention device of construction machinery.
In the hydraulic-pump driving system controlling device of FIG. 7, a variable-capacity hydraulic pump (main pump) 2 which is driven by an engine (internal-combustion engine) 1, and a pilot pump 3 are provided. A discharge outlet of the variable capacity hydraulic pump 2 communicates with a control valve 4 which controls supply and exhaust of hydraulic pressure from the variable capacity hydraulic pump 2 to a hydraulic actuator which is not illustrated.
In the hydraulic-pump driving system controlling device of FIG. 7, pilot ports 4a which are provided at both ends of the control valve 4 respectively communicate with a pilot-pressure discharge outlet of a control lever 6 via a pilot-pressure introducing line 5. A pilot pressure from the pilot pump 3 is introduced into the control lever 6 via the line which is not illustrated, and the introduced pressure is used as a pilot pressure to operate the control valve 4.
Moreover, the discharge outlet of the variable capacity hydraulic pump 2 communicates with a hydraulic pressure inlet of a regulator (discharge-quantity control unit) 7 via a line 13. The variable capacity hydraulic pump 2 supplies a discharge pressure to the regulator 7 to decrease the discharge quantity in proportion to an increase in the discharge pressure. Thus, the variable capacity hydraulic pump 2 is operated by performing a constant-torque control (or constant-horsepower control) which controls the input torque uniformly so that the input torque may not exceed an engine torque.
Moreover, the variable capacity hydraulic pump 2 is operated by performing a flow control which increases or decreases the discharge quantity in accordance with the control input of the control lever 6. FIG. 8 shows the constant-horsepower control which is performed in the hydraulic-pump driving system controlling device of FIG. 7, and the constant-horsepower curves (in H mode and L-mode) are indicated.
If a hydraulic excavator including a hydraulic actuator is considered as a typical construction machinery, various construction operations, including heavy-load digging, light-load digging, finishing, etc. are performed using the hydraulic excavator.
In order to control the input torque of the variable capacity hydraulic pump 2 so that an optimal input torque for one of the various construction operations may be selected, the hydraulic-pump driving system controlling device of FIG. 7 is provided with a mode selector switch 8, a controller (control unit) 9, and an electromagnetic inverse-proportion valve (input torque control unit) 10. The mode selector switch 8 outputs an external signal. The controller 9 receives this external signal from the mode selector switch 8 and outputs a torque setting signal. The electromagnetic inverse-proportion valve 10 receives this torque setting signal from the controller 9 and outputs a secondary pressure Pf.
The mode selector switch 8, the controller 9, and the electromagnetic inverse-proportion valve 10 mentioned above constitute an operation-mode selector circuit. The secondary pressure Pf from the electromagnetic inverse-proportion valve 10 is supplied to the regulator 7, and as shown in FIG. 8, the input torque of the variable capacity hydraulic pump 2 is changed between Tmax and Tmin, and the input torque is set to an input torque value between Tmax and Tmin according to the level of the external signal from the mode selector switch 8.
FIG. 9 is a time chart for explaining the respective characteristics of the parts of the hydraulic-pump driving system controlling device of FIG. 7 when usual digging is performed using a hydraulic excavator as construction machinery and the constant-horsepower control is set in the H mode.
If sudden actuation of the control lever 6 is performed as shown in FIGS. 9 (a) and (b), the discharge quantity Q of the variable capacity hydraulic pump 2 begins to increase. Simultaneously, in order to operate the hydraulic actuator, starting pressure occurs, and the discharge pressure P of the variable capacity hydraulic pump 2 increases rapidly to P1 (see FIG. 9 (c)).
When the constant-horsepower control is set in the H mode, in order to control uniformly the input torque of the variable capacity hydraulic pump 2, the secondary pressure Pf of the electromagnetic inverse-proportion valve 10 is set to the predetermined value Pf1 (see FIG. 9 (f)).
Since the constant-horsepower control set in the H mode cannot respond to the sudden rise to the discharge pressure P1 at this time, while the discharge pressure P of the variable capacity hydraulic pump 2 increases quickly, the input torque T of the variable capacity hydraulic pump 2 exceeds the torque when the engine speed N is at the nominal-speed N0, and it is set to T1 (see FIG. 9 (d)).
As a result, the engine speed N of the engine 1 falls to the engine speed N1 at which the torque is balanced, and the pump discharge quantity Q temporarily falls with this lowering (lag down) of the engine speed (see FIGS. 9 (e) and (b)).
Once the hydraulic actuator operates, the sliding state of the respective parts changes from static friction to dynamic friction and the pump discharge pressure P falls to P2. The input torque T of the variable capacity hydraulic pump 2 also falls to Tmax and the pump discharge quantity Q increases to Q1, thereby returning to the control state of the constant-horsepower control.
However, while the engine speed is falling, controlling the engine 1 to increase the fuel injection quantity is performed in order to return the engine-speed N to the nominal speed N0. As shown in FIGS. 9 (e) and (g), the control to increase the lowered engine-speed N back to the nominal speed N0 is performed by increasing the fuel injection quantity q of the engine 1 from q1 to q2 at the time of lowering of the engine speed. By increasing the fuel injection quantity q from q1 to q2, the fuel injection quantity equivalent to the shaded portion F indicated in FIG. 9 (g) will be a cause of increase in the fuel consumption of the engine 1.
For example, Japanese Laid-Open Patent Application No. 2005-76670 discloses an engine lag-down prevention device of construction machinery which is known as a conventional over-loading prevention device of construction machinery. This engine lag-down prevention device includes a main pump which is driven by an engine, a torque control valve which adjusts a maximum pump torque of the main pump, a hydraulic actuator which is driven by a hydraulic pressure supplied from the main pump, and an operation device which operates the hydraulic actuator.
Moreover, in the engine lag-down prevention device, a torque control unit is arranged. This torque control unit is arranged to control the torque control valve to gradually increase the hydraulic pump torque based on a predetermined torque increasing rate with the progress of time from the end of a predetermined torque holding time for which the low pump torque is held, immediately after the operation device is operated from the non-operating state.
Since the engine lag-down prevention device of Japanese Laid-Open Patent Application No. 2005-76670 is arranged so that the hydraulic-pump torque is increased gradually by the torque control unit, the load acting on the engine can be reduced even after the end of the predetermined torque holding time. Accordingly, the engine lag-down after the end of the predetermined torque holding time can be reduced to a small amount.
Patent Document 1: Japanese Laid-Open Patent Application No. 2005-76670